Hydraulic control system having accumulator

ABSTRACT

A hydraulic control system for preventing shortage of lubricant and coolant during charging of an accumulator is disclosed. The hydraulic control system is comprised of: a hydraulic pump; an accumulator for storing fluid from the pump; and a constant delivery site which is connected to the pump in parallel with the accumulator, and which requires hydraulic fluid constantly during activation of the hydraulic pump. A pressure level of the fluid discharged from the pump is regulated to a level possible to be delivered to the accumulator, and the fluid resulting from the pressure regulation is delivered to the constant delivery site. Meanwhile, a pressure higher than that required by the constant delivery site is stored in the accumulator. The hydraulic control system is characterized by a fluid increasing means that delivers additional fluid to the constant delivery site together with the fluid resulting from said pressure regulation, when delivering the fluid discharged from the pump to the accumulator.

TECHNICAL FIELD

This invention relates to a hydraulic control system for delivering hydraulic fluid to various types of actuators and lubrication points, while storing hydraulic pressure in an accumulator.

BACKGROUND ART

Various kinds of actuating mechanisms are actuated hydraulically, and hydraulic fluid is partially used for cooling and lubrication. Usually, the hydraulic pressure is established by a hydraulic pump driven by an engine or a motor. That is, cessation of engine or motor operation will result in a loss of hydraulic pressure. In order to maintain the hydraulic pressure to continue an operation in progress in case of stopping the hydraulic pump, an accumulator is used in a conventional hydraulic system. For example, conventional hydraulic control devices are disclosed in Japanese Patent Laid-Open No. 2004-84928 and International Publication No. 2010/021218.

The control devices disclosed in those publications are used to control a transmission. According to the teachings of Japanese Patent Laid-Open No. 2004-84928, a selection valve is arranged on an oil passage connected with a discharge port of a motor-driven oil pump for the purpose of selectively delivering hydraulic fluid to an accumulator and to a lubrication system. Meanwhile, the control device disclosed in International Publication No. 2010/021218 is applied to continuously variable transmissions. According to the teachings of International Publication No. 2010/021218, hydraulic pressure established by an oil pump is delivered to both a high pressure section including the CVT and a clutch, and a low pressure section including a torque convertor and lubrication points. In addition, in order to achieve a required transmission torque capacity of the transmission mechanism during a stop event of the oil pump, an accumulator for storing the hydraulic fluid to be delivered to the CVT and forward clutch is arranged in the high pressure section.

As discussed in the previously cited prior art documents, the hydraulic fluid is usually delivered by a single pump, to both actuator and the lubrication and cooling spots. That is, in the hydraulic control device taught by Japanese Patent Laid-Open No. 2004-84928, the hydraulic fluid will not be delivered to the lubrication and cooling spots when the selection valve is in a position to deliver the hydraulic fluid to the accumulator. In this situation, therefore, it is difficult to sufficiently lubricate the lubrication spot and cool the cooling spot due to lack of hydraulic fluid.

In turn, in the hydraulic control device taught by International Publication No. 2010/021218, the high pressure section including the accumulator and the low pressure section including lubrication spots are connected in parallel. Therefore, the hydraulic fluid cannot be delivered sufficiently to the lubrication spots through the regulator valve when charging the accumulator by delivering high pressured fluid thereto. In addition, an engine-driven mechanical oil pump incapable of varying a discharge rate thereof is used in the hydraulic control system for automobiles in most cases. As such, the lubrication spots will not be lubricated sufficiently and the cooling spots will not be cooled sufficiently during accumulator charging.

DISCLOSURE OF THE INVENTION

In order to solve the foregoing technical problems, it is an object of this invention to provide a hydraulic control system, which provides a sufficient lubrication and cooling of lubrication points during a fluid charge of a hydraulic accumulator.

The hydraulic control system of the present invention is comprised of a hydraulic pump, an accumulator that stores hydraulic fluid discharged from the hydraulic pump, and a constant delivery site connected to the hydraulic pump in parallel with the accumulator, that requires hydraulic fluid constantly during activation of the hydraulic pump. According to the present invention, the hydraulic fluid resulting from regulating a pressure level of the hydraulic fluid discharged from the hydraulic pump to a level possible to be delivered to the accumulator side is delivered to the constant delivery site, and the hydraulic pressure higher than the pressure required by the constant delivery site is stored in the accumulator. In order to solve the above-explained technical problems, the hydraulic control system of the present invention is provided with a fluid increasing means that delivers additional fluid to the constant delivery site together with the hydraulic fluid resulting from said pressure regulation, when delivering the hydraulic fluid discharged from the hydraulic pump to the accumulator.

The fluid increasing means is comprised of: a high pressure passage for delivering the hydraulic fluid discharged from the hydraulic pump to the accumulator; and a flow volume increasing device that is activated by the hydraulic fluid delivered from the high pressure passage to increase a flow volume of the hydraulic fluid being delivered to the constant delivery site.

The hydraulic control system of the present invention further comprises a constant supply passage for delivering the hydraulic fluid resulting from said pressure regulation to the constant delivery site; and a bypass passage connecting the high pressure passage with the constant supply passage. In addition, a jet pump may be used as the flow volume increasing device. In this case, the jet pump is disposed on the bypass passage. Specifically, the jet pump is adapted to introduce fluid therein from a reservoir by a negative pressure resulting from spraying the hydraulic fluid flowing through the bypass passage into the jet pump.

The hydraulic control system further comprises a shutoff device that blocks off the flow of the hydraulic fluid to the flow volume increasing device when the hydraulic fluid is not delivered from the hydraulic pump to the accumulator through the high pressure passage.

The shutoff device includes: a first on/off valve that is disposed on the high pressure passage between the hydraulic pump and a branching point of the bypass passage, and that is opened to deliver the hydraulic fluid to the accumulator; and a second on/off valve that is disposed on the bypass passage between an outlet of the flow volume increasing device and the constant supply passage, and that is opened to deliver the hydraulic fluid to the constant supply passage.

A check valve, that is adapted to be opened elastically when the pressure of the hydraulic fluid flowing toward the accumulator exceeds a predetermined opening level, and that is adapted to block the hydraulic fluid flowing backwardly from the accumulator, may be used as the first on/off valve.

Meanwhile, a check valve that is opened by the hydraulic fluid delivered to the constant supply passage may be used as the second on/off valve.

The hydraulic control system of the present invention further comprises an engine that drives the hydraulic pump, and an electric pump that is driven by a motor to discharge the hydraulic fluid. In addition, the fluid increasing means includes a switching valve that is switched to deliver the hydraulic fluid discharged from the electric pump to the constant delivery site when the hydraulic fluid is being delivered from the hydraulic pump to the accumulator.

The switching valve includes a valve that provides communication between the electric pump and the constant delivery site when the hydraulic fluid discharged from the hydraulic pump is regulated to a pressure level possible to be delivered to the accumulator side, and that provides communication between the electric pump and the accumulator side when the hydraulic fluid discharged from the hydraulic pump is regulated to a pressure level lower than the level at which the hydraulic fluid is allowed to be delivered toward the accumulator side.

The hydraulic control system of the present invention further comprises a deficiency calculation means that calculates a deficit flow quantity of the hydraulic fluid being delivered from the hydraulic pump to the constant delivery site when delivering the hydraulic fluid discharged from the hydraulic pump to the accumulator side; and an electric pump control means that controls the electric pump based on the deficit flow quantity of the hydraulic fluid calculated by the deficiency calculation means.

The electric pump includes an oil pump that is adapted to discharge the hydraulic fluid in accordance with a rotational speed thereof. Meanwhile, the electric pump control means includes a means that controls a rotational speed of the electric pump or the motor in a manner such that the electric pump is allowed to discharge the hydraulic fluid in an amount to compensate for the deficiency of the hydraulic fluid calculated by the deficiency calculation means.

The hydraulic control system of the present invention further comprises a fluid coupling having a lock-up clutch that is engaged and disengaged hydraulically, and a control valve that creates a hydraulic pressure for controlling the lock-up clutch. In addition, the fluid increasing means comprises a flow volume increasing device that is driven by the hydraulic fluid delivered to the lock-up clutch to increase the flow volume of the hydraulic fluid being delivered to the constant delivery site.

The flow volume increasing device includes a jet pump, that is disposed between a passage for delivering the hydraulic fluid to the lock-up clutch and the passage for delivering the hydraulic fluid to the constant delivery site, and that is adapted to introduce the fluid therein from a reservoir by a negative pressure resulting from spraying the hydraulic fluid being delivered to the lock-up clutch into the jet pump.

The hydraulic control system of the present invention further comprises a switching valve, that is disposed before an inlet of the jet pump, and that is opened to allow the hydraulic fluid to be delivered to the jet pump when the pressure level of the hydraulic fluid being delivered to the accumulator side is high, and closed to block the flow of the hydraulic fluid toward the jet pump when the pressure level of the hydraulic fluid being delivered to the accumulator side is low.

The hydraulic control system of the present invention further comprises a pressure regulator valve that regulates a hydraulic pressure established by the hydraulic pump. The hydraulic fluid drained from the pressure regulator valve as a result of regulating the hydraulic pressure is delivered to the constant delivery site.

The constant delivery site includes at least any one of a lubrication point and a cooling spot cooled by oil.

The hydraulic control system of the present invention further comprises a pressure regulator valve that regulates a hydraulic pressure established by the hydraulic pump while outputting a signal pressure in accordance with a pressure regulating level. The hydraulic fluid drained from the pressure regulator valve as a result of regulating the hydraulic pressure is delivered to the constant delivery site, and the switching valve is switched by the signal pressure.

Thus, according to the present invention, the pressure level of the hydraulic fluid delivered to the accumulator is higher than that of the fluid delivered to the constant delivery site such as the lubrication site, and the hydraulic pump discharges the hydraulic fluid to be delivered to both the accumulator and the constant delivery site. Therefore, when the hydraulic pressure is regulated to the higher level for the purpose of charging the accumulator, a quantity of the fluid drained as a result of the pressure regulation is reduced. However, when charging the accumulator, the fluid increasing means is activated to deliver additional fluid to the constant delivery site together with the hydraulic fluid discharged from the hydraulic pump. Therefore, a shortage of the fluid at the constant delivery site will not be caused even when charging the accumulator.

According to the present invention, the flow volume increasing device, that is, the jet pump driven by the relatively high pressured fluid delivered to the accumulator or the fluid delivered to the lock-up clutch to introduce the fluid in the reservoir serves as the fluid increasing means. Therefore, it is unnecessary to use an additional hydraulic source to increase the flow volume.

The hydraulic control system of the present invention is configured to avoid a fluid shortage at the constant delivery site when charging the accumulator. Therefore, the shutoff device prevents an increase in the hydraulic fluid to be delivered to the constant delivery site when the accumulator is not being charged. Therefore, the hydraulic fluid will not be delivered to the constant delivery site excessively so that a power loss can be reduced.

According to the present invention, the electric pump for delivering the hydraulic fluid to the accumulator side may also be used to increase the flow volume of the hydraulic fluid to be delivered to the constant delivery site. In addition, a discharge amount of the electric pump may be calculated based on a rotational speed thereof so that the electric pump can be controlled to discharge the fluid in an amount to compensate for the deficiency of the fluid at the constant delivery site. Therefore, the hydraulic fluid can be delivered to the constant delivery site in an exact required amount so that the electric pump will not be driven unnecessarily.

Thus, according to the present invention, not only a shortage of the lubricating fluid but also a shortage of cooling fluid can be prevented when charging the accumulator.

As described, the jet pump may be used as the flow volume increasing device to increase the fluid to be delivered to the constant delivery site. That is, mechanical parts to be rotated and moved linearly are not used in the jet pump. Therefore, in addition to the foregoing advantages, a structure of the hydraulic control system can be simplified and this makes maintenance of the hydraulic control system easier.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a block diagram schematically illustrating an exemplary hydraulic control system of the invention.

FIG. 2 is a block diagram schematically illustrating another exemplary hydraulic control system of the invention.

FIG. 3 is a block diagram schematically illustrating yet another exemplary hydraulic control system of the invention.

FIG. 4 is a block diagram schematically illustrating still another exemplary hydraulic control system of the invention.

FIG. 5 is a flowchart illustrating an exemplary control of an electric oil pump shown in FIG. 4.

BEST MODE FOR CARRYING OUT THE INVENTION

Referring now to FIG. 1, there is shown an example of applying the hydraulic control system of the invention to a continuously variable transmission used in an automobile. The continuously variable transmission 1 is a conventional belt-driven CVT comprising a drive pulley 2, a driven pulley 3 and a belt 4 mounted driveably on those pulleys 2 and 3 to transmit a torque therebetween. Each pulley 2 and 3 comprises a fixed sheave and a movable sheave being adapted for movement toward and away from the fixed sheave, and inwardly facing conical faces of the sheaves form a V-groove for holding the belt 4 therebetween. In order to move the movable sheaves axially, the pulleys 2, 3 are provided individually with hydraulic actuators 5, 6. In this preferred example, the movable sheave of the driven pulley 3 is moved axially in response to the application of pressurized hydraulic fluid to the actuator 6 thereby establishing a clamping pressure for clamping the belt 4 by the driven pulley 3. In turn, the movable sheave of the drive pulley 2 is moved axially in response to the application of pressurized hydraulic fluid to the actuator 5 thereby changing an effective diameter position of the belt 4 on the pulley 2 to vary the speed ratio.

C1 clutch 7 is disposed on either input or output side of the continuously variable transmission 1 to selectively transmit a drive torque. In this preferred example, a wet-type multiple disc clutch serves as the C1 clutch 7, and the transmitting torque capacity of the C1 clutch 7 varies in accordance with hydraulic pressure applied thereto. Specifically, the torque transmitting capacities of the continuously variable transmission 1 and the C1 clutch 7 are controlled hydraulically in such a way to transmit the torque to propel the vehicle. To this end, high-pressure hydraulic fluid is delivered to the actuators 5, 6 and C1 clutch 7 to increase the torque transmitting capacities. Accordingly, the continuously variable transmission 1, the hydraulic actuators 6, 7, the C1 clutch 7 and a (not shown) hydraulic chamber thereof form a high pressure section.

A conventional torque converter 8 having a lock-up clutch (not shown) is disposed in tandem with the continuously variable transmission 1. The torque converter 8 is adapted to multiply torque when a speed ratio between a pump impeller and a turbine runner is within a range smaller than a predetermined value, in other words, difference between rotational speeds of the pump impeller and the turbine runner is large. However, in contrast, the torque converter 8 merely serves as a fluid coupling without multiplying the torque when the speed ratio between the pump impeller and the turbine runner is within a range larger than the predetermined value, in other words, difference between rotational speeds of the pump impeller and the turbine runner is small. The lock-up clutch physically links a front cover integrated with the pump impeller and a hub integrated with the turbine runner via a friction plate.

A lock-up control valve 9 is provided to control hydraulic pressure applied to the lock-up clutch of the torque converter. Specifically, the control valve 9 alters a direction and a pressure of the hydraulic fluid delivered to the lock-up clutch to provide for a reciprocating motion of the lock-up clutch, thereby engaging the friction plate of the lock-up clutch selectively with the front cover. To this end, the control valve 9 is adapted to be actuated by relatively low pressure.

In the power train including the continuously variable transmission 1 and the torque converter 8, the hydraulic fluid is delivered to a number of frictional contact points and sliding members such as bearings for the purpose of cooling and lubricating those frictional contact points and sliding members. For these purposes, it is unnecessary to pressurize the hydraulic fluid but necessary to deliver required amount of fluid to the lubrication point 10. Accordingly, the lubrication point 10, the control valve 9, and the torque converter 8 form a low pressure section.

In the preferred example shown in FIG. 1, a mechanical hydraulic pump 12, which is driven by the engine 11 of the vehicle, is provided to deliver the hydraulic fluid to both the high pressure section and the low pressure section. In order to regulate pressure of the hydraulic fluid delivered from the mechanical hydraulic pump 12 to the low pressure section, a pressure regulator valve 13 is disposed in the low pressure section. For example, the pressure regulator valve 13 is adapted to regulate the pressure of the hydraulic fluid discharged from the mechanical hydraulic pump 12 to a relatively low pressure demanded by a feeding port of the control valve 9 to activate the lockup clutch. Specifically, the pressure regulator valve 13 is adapted to regulate the pressure of the hydraulic fluid delivered thereto to a target level by partially discharging the hydraulic fluid therefrom. The discharged hydraulic fluid (i.e., a drain pressure) is delivered to the lubrication point 10 through an oil passage 14. Accordingly, the lubrication point 10 corresponds to the constant delivery site of this invention to which the hydraulic fluid is delivered constantly, and the oil passage 14 serves as the constant supply passage of this invention.

An alternator 15 shown in FIG. 1 is also driven by the engine 11 to serve as an electricity generator. As the conventional alternators used in engine driven automobiles, the alternator 15 is adapted to supply electric power to electrical components while storing in a not shown battery. In the preferred example shown in FIG. 1, the electric power generated by the alternator 15 is supplied to a motor 17 to activate an electric pump 16. The main role of the electric pump 16 is to deliver the hydraulic fluid to the high pressure section. Therefore, a discharge outlet of the electric pump 16 is connected with an accumulator 20 through a passage 18, and a check valve 19 is disposed on the passage 18 between the electric pump 16 and the accumulator 20. In case discharge pressure of the electric pump 16 is lower than the pressure stored in the accumulator 20, the check valve 15 is closed to prevent backflow from the accumulator 20 to the electric pump 16.

In order to deliver the hydraulic fluid to the high pressure section from the electric pump 16 or the accumulator 20, the passage 18 is divided into three branches to be connected individually with the actuator 5 of the drive pulley 2, the actuator 6 of the driven pulley 3 and the C1 clutch 7. A solenoid valve is disposed on each branch for the purpose of controlling hydraulic pressures applied to the actuators 5, 6 and the C1 clutch 7. Specifically, a feeding solenoid valve DSP1 is disposed on the branch connected with the actuator 5 of the drive pulley 2, and the actuator 5 is also connected with a drain solenoid valve DSP2. For example, a groove width of the drive pulley 2 is narrowed (to carry out an upshifting) by opening the feeding solenoid valve DSP1 thereby delivering the hydraulic fluid to the actuator 5. In contrast, the groove width of the drive pulley 2 is widened (to carry out a downshifting) by opening the drain solenoid valve DSP2 thereby discharging the hydraulic fluid in the actuator 5 to an open area such as an oil reservoir. The speed ratio achieved by the solenoid valves DSP1 and DSP2 may be controlled based on a feedback and feedforward principles using a deviation between a target ratio and an actual ratio.

Deliveries of hydraulic fluid to the actuator 6 of the driven pulley 3 and to the C1 clutch 7 is controlled in a similar way. That is, in turn, a feeding solenoid valve DSS1 is disposed on the branch connected with the actuator 6 of the driven pulley 3, and the actuator 6 is also connected with a drain solenoid valve DSS2. Therefore, the clamping force (or pressure) of the driven pulley 3 to clamp the belt 4 is increased by opening the feeding solenoid valve DSS1 thereby delivering the hydraulic fluid to the actuator 6. In contrast, the clamping force of the driven pulley 3 is weakened by opening the drain solenoid valve DSS2 thereby discharging the hydraulic fluid in the actuator 6 to the open area such as the oil reservoir. The clamping force of the driven pulley 3 may also be controlled based on a feedback and feedforward principles on the basis of a drive demand which may be detected from, e.g., an opening degree of an accelerator.

Likewise, a feeding solenoid valve DSC1 is disposed on the branch connected with the C1 clutch 7, and the C1 clutch 7 is also connected with a drain solenoid valve DSC2. Therefore, the C1 clutch 7 is engaged by opening the feeding solenoid valve DSC1 thereby delivering the hydraulic fluid to the C1 clutch 7, and disengaged by opening the drain solenoid valve DSC2 thereby discharging the hydraulic fluid from the C1 clutch 7.

As described, the mechanical hydraulic pump 12 is activated by the engine 11, and a pressure of the hydraulic fluid discharged from the mechanical hydraulic pump 12 is regulated by the pressure regulator valve 13. In order to store the pressurized hydraulic fluid discharged from the mechanical hydraulic pump 12 in the accumulator 20, the outlet of the mechanical hydraulic pump 12 is connected with the accumulator 20 through a passage 21. Accordingly, the passage 21 serves as the high pressure passage of this invention, and a check valve 22 is disposed on the passage 21 to prevent backflow of the hydraulic fluid from the accumulator 20 toward the mechanical hydraulic pump 12. In addition, a region of the passage 21 containing the hydraulic fluid whose pressure is regulated by the pressure regulator valve 13, that is, a region of the passage 21 between the check valve 22 and the outlet of the mechanical hydraulic pump 12 is communicated with the passage 14 through a passage 23. Thus, the passage 23 serves as the bypass passage of this invention and to be called a bypass passage 23 hereinafter.

A flow volume increasing device 24 is disposed on the bypass passage 23 for the purpose of increasing a volume of the hydraulic fluid to be delivered to the lubrication point 10. In addition, an orifice 25 is formed on the bypass passage 23 before the flow volume increasing device 24 (i.e., upstream side of the flowing direction) to reduce the pressure of the hydraulic fluid flowing therethrough. The flow volume increasing device 24 is used to introduce the fluid therein from a reservoir 26 such as an oil pan utilizing fluid energy of the hydraulic fluid flowing therethrough and to deliver the introduced fluid to the lubrication point 10. For these applications, for example, a conventional jet pump (or an ejector pump) may be employed as the flow volume increasing device 24.

The flow volume increasing device 24 is comprised of a nozzle 27 having a decreased diameter for providing increased velocity of the fluid flowing therethrough, and a diffuser 28 having an increased outlet diameter. Therefore, the hydraulic fluid is sprayed from the nozzle 27 into the diffuser 28 at high velocity so that a negative pressure is created within an inlet of the diffuser 28. As a result of development of the negative pressure, the hydraulic fluid stored in the reservoir 26 will gravitate into the inlet of the diffuser 28 to be mixed with the hydraulic fluid being discharged from the nozzle 27. Therefore, the fluid in the reservoir 26 will also be delivered from the flow volume increasing device 24 to the lubrication point 10 in addition to the hydraulic fluid discharged from the nozzle 27.

When the engine is running e.g., to propel the vehicle, it is necessary to control the torque transmitting capacity, that is, the clamping pressure of the continuously variable transmission 1 in accordance with an output torque of the engine 11, and to control the speed ratio of the continuously variable transmission 1 in accordance with a vehicle speed, an opening degree of the accelerator and so on. Therefore, the hydraulic control system shown in FIG. 1 is required to establish hydraulic pressure to achieve desired torque transmitting capacity and speed ratio. To this end, the electric pump 16 is used mainly to generate the required pressure and the accumulator 20 may also be used in place of the electric pump 16. The relatively high-pressure fluid discharged from the electric pump 16 or the accumulator 20 is delivered to the continuously variable transmission 1 and the C1 clutch 7 through the passage 18. As described, the hydraulic pressures applied to the actuators 5,6 and the C1 clutch 7 are controlled by electrically opening and closing the feeding solenoid valves DSP1, DSS1, DSC1 and the drain solenoid valves DSP2, DSS2, DSC2. When the engine 11 is thus running, the transmission members such as the torque converter 8 and other gears are being rotated. In this situation, therefore, the mechanical hydraulic pump 12 being driven by the engine 11 is required to deliver the hydraulic fluid to the lubrication point 10 and the torque converter 8. Pressure of the hydraulic fluid discharged from the mechanical hydraulic pump 12 is regulated to a predetermined level by the pressure regulator valve 13 and delivered to the control valve 9 thereby actuating the lock-up clutch. The hydraulic fluid drained from the pressure regulator valve 13 as a result of thus regulating the fluid pressure is delivered to the lubrication point 10 through the passage 14 thereby lubricating the bearings, tooth flanks of the gears and so on.

At the same time, the hydraulic fluid discharged from the mechanical hydraulic pump 12 flows partially into the bypass passage 23 through the high pressure passage 21. As described, the hydraulic fluid flowing through the bypass passage 23 is sprayed from the nozzle 27 while crating the negative pressure to introduce the fluid in the reservoir 26 into the flow volume increasing device 24. Therefore, the fluid in the reservoir 26 is delivered to the lubrication point 10 together with the hydraulic fluid flowing through the bypass passage 23. That is, an amount of the hydraulic fluid delivered to the lubrication point 10 is increased. Additionally, a cooling spot may be connected to the oil passage 14 in parallel with the lubrication point 10 to be cooled by the hydraulic fluid. Those lubrication point 10 and the cooling spot have to be supplied with the hydraulic fluid constantly during running of the engine 11. According to this invention, however, not only the hydraulic fluid discharged from the mechanical hydraulic pump 12 but also the fluid in the reservoir 26 are delivered to the lubrication point 10 and the cooling spot during driving the mechanical hydraulic pump 12 by the engine 11. Therefore, ample amount of the hydraulic fluid can be delivered certainly to the lubrication point 10 and the cooling spot to avoid shortage of the lubrication fluid or cooling fluid.

Likewise, the accumulator 20 may also be supplied with ample amount of the hydraulic fluid. For example, provided that the pressure in the accumulator 20 drops, or when the pressure regulator valve 13 regulates the pressure of the hydraulic fluid to the higher level, the mechanical hydraulic pump 12 outputs the higher hydraulic pressure. In such cases, the check valve 22 is opened by the pressure of the hydraulic fluid delivered to the high pressure passage 21 so that the flow volume of the hydraulic fluid delivered from the mechanism hydraulic pump 12 to the accumulator 20 is increased. Instead, the flow volume of the hydraulic fluid delivered from the mechanical hydraulic pump 12 to the constant delivery site such as the lubrication point 10 and the torque converter 8 is relatively reduced. In this situation, however, the hydraulic fluid flowing through the high pressure passage 21 is partially delivered to the flow volume increasing device 24 through the bypass passage 23, and delivered to the lubrication point 10 together with the fluid introduced from the reservoir 26. Therefore, sufficient amount of the hydraulic fluid can be delivered to the lubrication point 10 also in this case to avoid shortage of the lubrication fluid.

Thus, in the example shown in FIG. 1, the flow volume of the hydraulic fluid delivered to the lubrication point 10 is increased utilizing the fluid flowing toward the accumulator 20 through the high pressure passage 21 to be stored in the accumulator 20. In this example, however, delivery of hydraulic fluid to the high pressure passage 21 is continued even when the hydraulic fluid is not necessary to be stored in the accumulator 20. This means that the volume of the hydraulic fluid delivered to the lubrication point 10 is increased regardless of implementation of pressure charge of the accumulator 20. However, the flow volume increasing device 24 is provided for the purpose of increasing the supply of the fluid to the lubrication point 10 on the occasion of charging the accumulator 20 by delivering the hydraulic fluid thereto. Therefore, it is preferable to use the flow volume increasing device 24 only when charging the accumulator 20. An example therefore is shown in FIG. 2.

In the preferred example shown in FIG. 2, a shutoff device is provided to block off the flow of the hydraulic fluid to the flow volume increasing device 24 when the hydraulic fluid is not being delivered from the mechanical pump 12 to the accumulator 20. The elements identical to those in the example shown in FIG. 1 are represented by the common reference numerals, and detailed explanation thereof will be omitted. Specifically, the shutoff device employed in the example shown in FIG. 2 is comprised of a valve for blocking off the flow of the hydraulic fluid from the mechanical hydraulic pump 12 to the flow volume increasing device 24, and a valve for blocking block off a backflow of the hydraulic fluid from the lubrication point 10 toward the flow volume increasing device 24. For example, not only an electrically controlled on-off valve but also a hydraulically controlled on-off valve may be used to serve as those valves. In the example shown in FIG. 2, specifically, a pressure storing check valve 29 is disposed on the high pressure passage 21 between the mechanical hydraulic pump 12 and the check valve 22 preventing backflow of the hydraulic fluid from the accumulator 20 or from the oil passage 18. The pressure storing check valve 29 is opened elastically by the hydraulic fluid substantially at the minimum pressure level required to be stored in the accumulator 20. That is, the pressure storing check valve 29 stays closed when the hydraulic pressure generated by the mechanical hydraulic pump 12 is not being stored in the accumulator 20. The bypass passage 23 is branched off from the high pressure passage 21 between the check valve 22 and the pressure storing check valve 29. Accordingly, when the hydraulic fluid discharged from the mechanical hydraulic pump 12 is not being stored in the accumulator 20, the hydraulic fluid will not enter into the bypass passage 23.

Meanwhile, a check valve 30 is disposed between the outlet of the flow volume increasing device 24 and the oil passage 14 for delivering the hydraulic fluid discharged from the pressure regulator valve 13 to the lubrication point 10. The check valve 30 is adapted to be opened when the discharge pressure of the flow volume increasing device 24 is higher than the pressure in the oil passage 14, and closed when the discharge pressure of the flow volume increasing device 24 is lower than the pressure in the oil passage 14. Therefore, when the flow volume increasing device 24 is not activated so that the discharge pressure thereof is relatively low, the hydraulic fluid will not flow backwardly from the lubrication point 10 or from the oil passage 14 toward the flow volume increasing device 24.

Thus, according to the preferred example shown in FIG. 2, if the discharge pressure of the mechanical hydraulic pump 12 will not be stored in the accumulator 20, the pressure level of the hydraulic fluid to be delivered to the high pressure passage 21 is regulated by the pressure regulator valve 13 to be lower than the opening pressure of the pressure storing check valve 29. In this situation, therefore, the hydraulic fluid discharged from the mechanical hydraulic pump 12 is blocked by the pressure storing check valve 29 to be prevented from flowing into the bypass passage 23. Consequently, the flow volume increasing device 24 will not be activated to increase the flow volume of the hydraulic fluid delivered to the lubrication point 10. In addition, the hydraulic fluid discharged from the mechanical hydraulic pump 12 is delivered entirely to the torque converter 8 and the pressure regulator valve 13. Therefore, shortage of the lubrication fluid at the lubrication point 10 will not be caused. Especially, according to the preferred example shown in FIG. 2, the check valve 30 prevents the hydraulic fluid from flowing backwardly into the inactivated flow volume increasing device 24 (i.e., jet pump). Therefore, the lubrication fluid will be delivered to the lubrication point 10 in full measure.

According to the foregoing examples, the hydraulic fluid delivered to the accumulator 20 side is utilized to establish an initial pressure for activating the jet pump serving as the flow volume increasing device 24 by delivering the hydraulic fluid to the inlet of the jet pump through the bypass passage 23. However, the present invention may be modified to use another pressure to activate the jet pump. To this end, the hydraulic control system shown in FIG. 2 is partially modified as shown in FIG. 3, and common reference numerals are allotted to the common elements in FIG. 3. Therefore, detailed explanation for the common elements will be omitted and the structure different from that of the example shown in FIG. 2 will be explained hereinafter.

In the preferred example shown in FIG. 3, a bypass passage 32 is additionally formed to connect the inlet of the jet pump serving as the flow volume increasing device 24 (i.e., the inlet of the nozzle 27) with an oil passage 31 for delivering the hydraulic fluid regulated by the control valve 9 to the torque converter 8. In addition, a switching valve 33 adapted to be opened and closed by a signal pressure is disposed on the bypass passage 32. In this example, the pressure regulator valve 13 outputs the signal pressure according to (or based on) a pressure level to regulate the hydraulic fluid. More specifically, the pressure regulator valve 13 outputs the signal pressure when regulating the pressure of the hydraulic fluid to a higher level to deliver the hydraulic fluid to the accumulator 20 for the purpose of charging the accumulator 20, in accordance with the pressure level thus raised.

For example, a spool valve may be used as the switching valve 33. In the spool valve, the signal pressure and an elastic force of a spring counteract each other across the spool. The elastic force of the spring is adjusted to be smaller than a thrust force of the signal pressure so that the spring is compressed by the signal pressure applied thereto. Therefore, when the signal pressure is applied to the switching valve 33, the spool is pushed by the signal pressure while compressing the spring thereby opening the switching valve 33. Consequently, the hydraulic fluid for controlling the lock-up clutch is allowed to be delivered to the jet pump. To the contrary, when the signal pressure is not applied to the switching valve 33, the spool is pushed by the spring to close the switching valve 33 thereby blocking the flow of the hydraulic fluid delivered from the control valve 9 toward the jet pump.

In the preferred example shown in FIG. 3, it is unnecessary to deliver the hydraulic fluid from the mechanical hydraulic pump 12 to the accumulator 20 provided that a sufficient amount of hydraulic fluid is stored in the accumulator 20. In this case, therefore, the pressure regulator valve 13 regulates the hydraulic fluid to a relatively low level so that an ample amount of the hydraulic fluid is drained from the pressure regulator valve 13 to be delivered to the lubrication point 10. In addition, the pressure regulator valve 13 will not output the signal pressure to the switching valve 33 in this situation so that the switching valve 33 stays closed. Therefore, the hydraulic fluid for controlling the lock-up clutch will not be distributed from the passage 31 to the jet pump serving as the flow volume increasing device 24. That is, the jet pump will not be activated to deliver the fluid to the lubrication point 10 excessively. In addition, unnecessary power consumption can be avoided.

In contrast, when the pressure in the accumulator 20 drops, that is, when the hydraulic fluid has to be delivered from the mechanical hydraulic pump 12 to the accumulator 20, the pressure regulator valve 13 regulates the hydraulic fluid to a relatively high level. In this situation, therefore, the drain rate of the pressure regulator valve 13 is reduced so that the volume of the fluid to be delivered to the lubrication point 10 is reduced. However, as described, the pressure regulator valve 13 outputs the signal pressure when regulates the hydraulic pressure to a high level, and the signal pressure thus outputted is distributed to the switching valve 33 to open the switching valve 33. In this situation, therefore, the hydraulic fluid for controlling the lock-up clutch will be delivered to the jet pump through the switching valve 33. The hydraulic fluid thus delivered to the jet pump is then sprayed from the nozzle 27 into the diffuser 28 while crating the negative pressure to introduce the fluid in the reservoir 26 into the jet pump, and the check valve 30 is opened by the oil from the jet pump to allow the hydraulic fluid to be delivered to the lubrication point 10 together with the fluid from the reservoir 26. That is, the flow volume of the hydraulic fluid delivered to the lubrication point 10 is increased.

Thus, according to the hydraulic control system of the present invention, the high pressure required to control the continuously variable transmission 1 etc. is established by the mechanical hydraulic pump 12 activated by the engine 11, and the high pressure thus established may also be stored in the accumulator 20. In addition, the constant delivery site such as the lubrication point 10 may also be supplied with the low pressure fluid constantly in an appropriate amount.

Now referring to FIG. 4 which illustrates an example of increasing the volume of the fluid delivered to the lubrication point 10 utilizing the electric pump 16. According to the preferred example shown in FIG. 4, the fluid in the reservoir 26 is delivered to the lubrication point 10 by the electric pump 16 instead of the jet pump. In this example, a switching valve 34 for changing a destination of the fluid is disposed on the discharging side of the electric pump 16. The switching valve 34 is comprised of an input port communicated with the electric pump 16, an output port communicated with the accumulator 20 and the high pressure section through the check valve 19, and another output port communicated with the lubrication point 10 through a bypass passage 35. Therefore, the output port to be communicated with the input port is switched depending on whether the signal pressure is applied to the switching valve 34. For example, a spool valve may also be used as the switching valve 34. As described, in the spool valve, the signal pressure and an elastic force of a spring counteract each other across the spool. The elastic force of the spring is adjusted to be smaller than a thrust force of the signal pressure also in this example. For example, when the signal pressure is applied to the switching valve 34, the spool is pushed by the signal pressure while compressing the spring to connect the electric pump 16 with the bypass passage 35. In contrast, when the signal pressure is not applied to the switching valve 34, the spool is pushed by the spring to connect the electric pump 16 with the accumulator 20.

According to the preferred example shown in FIG. 4, the pressure regulator valve 13 is also adapted to output the signal pressure, and the signal pressure outputted from the pressure regulator valve 13 is distributed to the switching valve 34. Specifically, the pressure regulator valve 13 outputs the signal pressure to the switching valve 34 when regulating the hydraulic fluid to the high pressure level to charge the accumulator 20 with the hydraulic fluid, and the switching valve 34 is switched by the signal pressure thus outputted in a manner that the electric pump 16 is communicated with the lubrication point 10 through a bypass passage 35. The remaining configurations are identical to those of the examples shown in FIGS. 1 to 3. In FIG. 4, the elements identical to those in the examples shown in FIGS. 1 to 3 are represented by the common reference numerals, and detailed explanation thereof will be omitted.

The electric pump 16 is adapted to increase a discharge amount or a discharge pressure thereof in accordance with an increase in a rotational speed thereof or a rotational speed of the motor 17. For example, when charging the accumulator 20, that is, when delivering the hydraulic fluid to the high pressure section, the electric pump 16 or the motor 17 is driven at a speed possible to generate the pressure demanded by the accumulator 20 or the high pressure section which is known in advance. In contrast, the constant delivery site, that is, the lubrication point 10 requires the hydraulic fluid at lower pressure than the pressure to be stored in the accumulator 20, and the electric pump 16 is required to discharge merely a small amount of hydraulic fluid to increase the flow volume of the hydraulic fluid to be delivered to the lubrication point 10. To this end, the electric pump 16 or the motor 17 is controlled in a different manner to supplement the fluid being delivered to the lubrication point 10, in comparison with the normal control for delivering the fluid to the accumulator 20 or the high pressure section by the electric pump 16 or the motor 17.

FIG. 5 shows the control example for increasing the fluid to be delivered to the lubrication point 10 using the electric pump 16 or the motor 17. First of all, it is determined whether the command pressure Psol applied to the control solenoid (SOL) of the pressure regulator valve 13 is higher than the lowest signal pressure Pacc_min to activate the control solenoid to deliver the hydraulic fluid to the accumulator 20 (at step S1). Specifically, at step S1, it is determined whether the pressure level of the hydraulic fluid discharged from the mechanical hydraulic pump 12 being driven by the engine 11 is raised to the high pressure level to be delivered to the accumulator 20, that is, whether the accumulator 20 is being charged. If the answer of step S1 is NO, it is unnecessary to increase the flow volume of the hydraulic fluid being delivered to the lubrication point 10. In this case, therefore, the normal control of the electric oil pump 16 is carried out (at step S2). In contrast, if the command pressure Psol applied to the pressure regulator valve 13 is higher than the lowest signal pressure Pacc_min so that the answer of step S1 is YES, a deficit flow quantity Qlub is calculated (at step S3).

Specifically, the deficit flow quantity Qlub is calculated based on a rotational speed of the engine 11 and a pressure required to charge the accumulator 20. Since the mechanical hydraulic pump 12 is driven by the engine 11, a discharge rate of the mechanical hydraulic pump 12 can be calculated based on the rotational speed of the engine 11. Meanwhile, a delivery rate of the hydraulic fluid to the accumulator 20 can be calculated based on the pressure required to charge the accumulator 20. Accordingly, a volume of the hydraulic fluid delivered to the lubrication point 10 and the torque converter 8 can be calculated based on a difference between the discharge rate of the mechanical hydraulic pump 12 and the delivery rate of the hydraulic fluid to the accumulator 20. In addition, a quantity of the fluid required to be delivered to the lubrication point 10 is known in advance. Therefore, the deficit flow quantity Qlub can be calculated based on those values.

Then, a required rotational speed of the electric pump 16 is calculated based on the deficit flow quantity Qlub thus obtained (at step S4). A discharge rate of the electric pump 16 per rotation is governed by its design. Therefore, the required rotational speed of the electric pump 16 can be calculated based on the deficit flow quantity Qlub. Thereafter, a drive command of the motor 17 is outputted to achieve the required rotational speed of the electric pump 16 (at step S5).

According to the examples of the hydraulic control system shown in FIGS. 3 and 4, when the hydraulic fluid is not being delivered to the accumulator 20, the hydraulic fluid is regulated by the pressure regulator valve 13 to the relatively low pressure level so that large quantity of the hydraulic fluid is drained from the pressure regulator valve 13. Therefore, the fluid can be delivered sufficiently to the lubrication point 10. In addition, the pressure regulator valve 13 will not output the signal pressure to the switching valve 34 in this situation so that the switching valve 34 keeps connecting the electric pump 16 with the accumulator 20 connected with the high pressure section. Further, when the commend pressure applied to the pressure regulator valve 13 is thus relatively low, the answer of the judgment step S1 in FIG. 5 will be NO. In this case, therefore, the normal control is carried out to keep the electric pump 16 stopping unless the accumulator 20 and the high pressure section require hydraulic pressure. In other words, the electric pump 16 and the motor 17 will not be driven unnecessarily so that the power will not be wasted.

Then, when the pressure in the accumulator 20 drops and the hydraulic fluid is therefore required to be delivered from the mechanical hydraulic pump 12 to the accumulator 20, the pressure regulator valve 13 raises the pressure level and the drain volume of the pressure regulator valve 13 is reduced. As a result, the flow volume of the hydraulic fluid delivered to the lubrication point 10 is reduced. In this situation, however, the pressure regulator valve 13 outputs the signal pressure to the switching valve 34 in consequence of thus raising the pressure regulating level. As a result, the switching valve 34 is switched to connect the electric pump 16 with the lubrication point 10 through the bypass passage 35. At the same time, since the command pressure applied to the pressure regulator valve 13 is increased, the answer of the judgment step S1 in FIG. 5 will be YES. As a result, the electric pump 16 is driven to compensate for the deficiency of the fluid being delivered to the lubrication point 10, and the hydraulic fluid discharged from the electric pump 16 is additionally delivered to the lubrication point 10. Thus, the flow volume of the fluid to be delivered to the lubrication point 10 is increased to avoid shortage of the lubrication fluid.

Here will be explained a relation between the forgoing examples and the present invention. The jet pump, the electric pump 16, and the bypass passages 23, 32, 35 correspond to the fluid increasing means of the invention. The jet pump and the electric pump 16 also correspond to the flow volume increasing device of the invention. The pressure storing check valve 29 and the check valve 30 shown in FIG. 2 correspond to the shutoff device of the invention. The pressure storing check valve 29 also corresponds to the first on/off valve and the check valve 30 also corresponds to the second on/off valve of the invention. The electronic control unit for carrying out the controls of steps S3, S4 and S5 corresponds to the deficiency calculation means and the electric pump control means of the invention.

Although the hydraulic control system of the present invention is applied to the continuously variable transmission 1 in the foregoing examples, it is understood that the invention is not limited by the exact construction or method illustrated and described above. For instance, the present invention may also be applied to hydraulic control systems of other kind of transmission and machineries. That is, the present invention may be applied to any kind of hydraulic systems configured to deliver hydraulic fluid to both accumulator and delivery site from a common hydraulic pump. In addition, any kinds of pumps adapted to increase a flow volume of the fluid to be delivered to the lubricating point when charging the accumulator may also be used as the flow volume increasing device instead of the above-explained jet pump, ejector pump and electric pump. Further, the present invention be modified to deliver the fluid continuously to cooling spots instead of lubrication spots. Lastly, according to the present invention, a fluid coupling having a lock-up clutch may also be employed instead of the torque converter. 

1. A hydraulic control system, comprising: a hydraulic pump; an accumulator that stores hydraulic fluid discharged from the hydraulic pump; a constant delivery site connected to the hydraulic pump in parallel with the accumulator, that requires hydraulic fluid constantly during activation of the hydraulic pump; wherein the hydraulic fluid resulting from regulating a pressure level of the hydraulic fluid discharged from the hydraulic pump to a level possible to be delivered to the accumulator side is delivered to the constant delivery site; and wherein the hydraulic pressure higher than the pressure required by the constant delivery site is stored in the accumulator; the hydraulic control system further comprising: a fluid increasing means that delivers additional fluid to the constant delivery site together with the hydraulic fluid resulting from said pressure regulation, when delivering the hydraulic fluid discharged from the hydraulic pump to the accumulator.
 2. The hydraulic control system as claimed in claim 1, wherein the fluid increasing means comprises: a high pressure passage for delivering the hydraulic fluid discharged from the hydraulic pump to the accumulator; and a flow volume increasing device that is activated by the hydraulic fluid delivered from the high pressure passage to increase a flow volume of the hydraulic fluid being delivered to the constant delivery site.
 3. The hydraulic control system as claimed in claim 2, further comprising: a constant supply passage for delivering the hydraulic fluid resulting from said pressure regulation to the constant delivery site; and a bypass passage connecting the high pressure passage with the constant supply passage; wherein the flow volume increasing device includes a jet pump disposed on the bypass passage, that is adapted to introduce fluid therein from a reservoir by a negative pressure resulting from spraying the hydraulic fluid flowing through the bypass passage into the flow volume increasing device.
 4. The hydraulic control system as claimed in claim 2, further comprising: a shutoff device that blocks off the flow of the hydraulic fluid to the flow volume increasing device when the hydraulic fluid is not delivered from the hydraulic pump to the accumulator through the high pressure passage.
 5. The hydraulic control system as claimed in claim 4, wherein the shutoff device includes: a first on/off valve that is disposed on the high pressure passage between the hydraulic pump and a branching point of the bypass passage, and that is opened to deliver the hydraulic fluid to the accumulator; and a second on/off valve that is disposed on the bypass passage between an outlet of the flow volume increasing device and the constant supply passage, and that is opened to deliver the hydraulic fluid to the constant supply passage.
 6. The hydraulic control system as claimed in claim 5, wherein the first on/off valve includes a check valve that is adapted to be opened elastically when the pressure of the hydraulic fluid flowing toward the accumulator exceeds a predetermined opening level, and that is adapted to block the hydraulic fluid flowing backwardly from the accumulator.
 7. The hydraulic control system as claimed in claim 5, wherein the second on/off valve includes a check valve that is opened by the hydraulic fluid delivered to the constant supply passage.
 8. The hydraulic control system as claimed in claim 1, further comprising: an engine that drives the hydraulic pump; and an electric pump that is driven by a motor to discharge the hydraulic fluid; wherein the fluid increasing means includes a switching valve that is switched to deliver the hydraulic fluid discharged from the electric pump to the constant delivery site when the hydraulic fluid is being delivered from the hydraulic pump to the accumulator.
 9. The hydraulic control system as claimed in claim 8, wherein the switching valve includes a valve that provides communication between the electric pump and the constant delivery site when the hydraulic fluid discharged from the hydraulic pump is regulated to a pressure level possible to be delivered to the accumulator side, and that provides communication between the electric pump and the accumulator side when the hydraulic fluid discharged from the hydraulic pump is regulated to a pressure level lower than the level at which the hydraulic fluid is allowed to be delivered toward the accumulator side.
 10. The hydraulic control system as claimed in claim 8, further comprising: a deficiency calculation means that calculates a deficit flow quantity of the hydraulic fluid being delivered from the hydraulic pump to the constant delivery site when delivering the hydraulic fluid discharged from the hydraulic pump to the accumulator side; and an electric pump control means that controls the electric pump based on the deficit flow quantity of the hydraulic fluid calculated by the deficiency calculation means.
 11. The hydraulic control system as claimed in claim 10, wherein the electric pump includes an oil pump that is adapted to discharge the hydraulic fluid in accordance with a rotational speed thereof; and the electric pump control means includes a means that controls a rotational speed of the electric pump or the motor in a manner such that the electric pump is allowed to discharge the hydraulic fluid in an amount to compensate for the deficiency of the hydraulic fluid calculated by the deficiency calculation means.
 12. The hydraulic control system as claimed in claim 1, further comprising: a fluid coupling having a lock-up clutch that is engaged and disengaged hydraulically; and a control valve that creates a hydraulic pressure for controlling the lock-up clutch; wherein the fluid increasing means comprises a flow volume increasing device that is driven by the hydraulic fluid delivered to the lock-up clutch to increase the flow volume of the hydraulic fluid being delivered to the constant delivery site.
 13. The hydraulic control system as claimed in claim 12, wherein the flow volume increasing device includes a jet pump, that is disposed between a passage for delivering the hydraulic fluid to the lock-up clutch and the passage for delivering the hydraulic fluid to the constant delivery site, and that is adapted to introduce the fluid therein from a reservoir by a negative pressure resulting from spraying the hydraulic fluid being delivered to the lock-up clutch into the jet pump.
 14. The hydraulic control system as claimed in claim 13, further comprising: a switching valve, that is disposed before an inlet of the jet pump, and that is opened to allow the hydraulic fluid to be delivered to the jet pump when the pressure level of the hydraulic fluid being delivered to the accumulator side is high, and closed to block the flow of the hydraulic fluid toward the jet pump when the pressure level of the hydraulic fluid being delivered to the accumulator side is low.
 15. The hydraulic control system as claimed in claim 1, further comprising: a pressure regulator valve that regulates a hydraulic pressure established by the hydraulic pump; and wherein the hydraulic fluid drained from the pressure regulator valve as a result of regulating the hydraulic pressure is delivered to the constant delivery site.
 16. The hydraulic control system as claimed in claim 1, wherein the constant delivery site includes at least any one of a lubrication point and a cooling spot cooled by oil.
 17. The hydraulic control system as claimed in claim 7, further comprising: a pressure regulator valve that regulates a hydraulic pressure established by the hydraulic pump while outputting a signal pressure in accordance with a pressure regulating level; wherein the hydraulic fluid drained from the pressure regulator valve as a result of regulating the hydraulic pressure is delivered to the constant delivery site; and wherein the switching valve includes a valve that is switched by the signal pressure.
 18. The hydraulic control system as claimed in claim 9, further comprising: a deficiency calculation means that calculates a deficit flow quantity of the hydraulic fluid being delivered from the hydraulic pump to the constant delivery site when delivering the hydraulic fluid discharged from the hydraulic pump to the accumulator side; and an electric pump control means that controls the electric pump based on the deficit flow quantity of the hydraulic fluid calculated by the deficiency calculation means.
 19. The hydraulic control system as claimed in claim 18, wherein the electric pump includes an oil pump that is adapted to discharge the hydraulic fluid in accordance with a rotational speed thereof; and the electric pump control means includes a means that controls a rotational speed of the electric pump or the motor in a manner such that the electric pump is allowed to discharge the hydraulic fluid in an amount to compensate for the deficiency of the hydraulic fluid calculated by the deficiency calculation means.
 20. The hydraulic control system as claimed in claim 14, further comprising: a pressure regulator valve that regulates a hydraulic pressure established by the hydraulic pump while outputting a signal pressure in accordance with a pressure regulating level; wherein the hydraulic fluid drained from the pressure regulator valve as a result of regulating the hydraulic pressure is delivered to the constant delivery site; and wherein the switching valve includes a valve that is switched by the signal pressure. 